The invention relates to a drag damper for a rotary-wing aircraft rotor comprising a hub and at least two blades each of which is connected to the hub by a corresponding connection device. The drag damper is intended for fitting between two elements, one of which is a flapping mass comprising one of the blades and the corresponding connection device, and the other element is one of the hub (in the conventional configuration) and an adjacent flapping mass on the rotor (in the so-called inter-blade configuration). In use, the drag damper will dampen the relative angular drag movements between said two elements, namely: in the first case (conventional configuration), the angular drag movements of said flapping mass relative to the hub, i.e. the angular deflections of the blade, and more generally of the corresponding flapping mass, about its drag axis, which is substantially parallel to the axis of rotation of the rotor; and in the second case (inter-blade configuration), the relative angular drag movements between said two adjacent flapping masses, i.e. the relative angular deflections of the corresponding two adjacent blades of the rotor about their drag axes, which are substantially parallel to the axis of rotation of the rotor.
The rotor is more particularly a helicopter main rotor subject to the instability phenomena known as ground resonance and air resonance, although a conventional tail rotor may also be equipped with drag dampers according to the invention.
On rotors of the hinged type, the device connecting a blade to the hub may be arranged as a means of securing the blade and hinging it to the hub, when the blade is connected by its root, possibly in the form of a fork, to the hub, or as a device which is substantially radial (relative to the rotor axis) generally termed a cuff, and fitted with yokes at the ends to be connected to the blade root on the one hand and on the other to means of securing and hinging, such as a spherical laminated stop, itself connecting it to the hub. On rotors of the semi-rigid type, this connecting device may be a flexible torsion arm, at the blade root, and surrounded by a torsionally rigid cuff integral with the blade root for controlling the blade in pitch, which is connected and hinged to the hub by this flexible torsion arm.
Numerous different embodiments of drag dampers are known, particularly dampers which are hydraulic, hydro-pneumatic, laminated with at least one layer of visco-elastic material stressed between two rigid fittings, or comprising combinations of these different means, these drag dampers comprising means of elastic return of defined stiffness and damping, when they are fitted to helicopter main rotors, to combat the resonance phenomena mentioned above.
It is a well-known practice to design helicopter rotor blades and therefore the corresponding flapping masses having a natural drag frequency, also termed first drag mode or natural drag mode, which is different from the nominal rotation frequency at which the rotor is designed to be driven.
More generally, to avoid in particular fatigue problems resulting from the dynamic stresses in the blades and the fuselage, and problems of vibration levels in the fuselage, it is essential to position correctly the natural frequencies of the blades in flapping, torsion and drag relative to the nominal rotation speed of the rotor and its harmonics (multiples).
This results from the fact that a helicopter rotor constitutes a powerful vibration generator. Because of the variable angles of incidence and speeds of rotor blades and also of helicopters, alternating loads of aerodynamic origin are developed particularly in the blades of rotors, and give rise in the latter to stresses as well as reactions on the attachments, particularly of the blades to the hubs. From this there result alternating loads and moments on the rotor heads, and the development of high vibration and stress levels in fuselages. The response of each blade, the stresses to which this blade is subjected and the loads which this blade transmits to the hub at the centre of the rotor are all the greater as at least one natural frequency of the blade (in drag, flapping and torsion) is close to the rotation frequency xcexa9 of the rotor or one of its harmonics nxcexa9 (where n is a whole number).
The dynamic characteristics of the rotor blades are therefore chosen to obtain suitable dynamic matching by ensuring that their natural vibration frequencies in flapping, drag and torsion are correctly positioned relative to the nominal rotation frequency xcexa9 of the rotor and its multiples n xcexa9, which is why it is necessary to observe certain simple rules for positioning the frequencies, and in particular two essential rules. The first rule is to avoid positioning a natural vibration frequency in flapping, drag or torsion on or very close to a harmonic of the rotation speed nxcexa9 (where nxe2x89xa71), and the second rule is to endeavour as far as possible to position only one of these three natural frequencies between two adjacent harmonics nxcexa9 and (n+1) xcexa9 of the rotation speed in order to avoid coupling. In addition to these two essential rules, it is imperative to follow recommendations proper to each type of deformation in flapping, drag or torsion.
Concerning the recommendations relating particularly to the drag is modes for hinged or semi-rigid (semi-hinged) rotors, the first drag mode (or natural drag frequency) is at the origin of ground resonance and air resonance problems due to coupling with modes of the helicopter structure.
On a rotor with blades hinged in drag, the angular frequency or pulsatance of the first drag mode is given by the expression:       ω    ⁢          xe2x80x83        ⁢    δ    =            Ω      ⁡              (                                            e              ·              M                        ⁢                          xe2x80x83                        ⁢            δ                                I            ⁢                          xe2x80x83                        ⁢            δ                          )                    1      /      2      
where e is the drag eccentricity of each blade, Mxcex4 is the static moment of the flapping mass (blade+device connecting it to the hub) relative to the hinge (drag axis) and Ixcex4 is the inertia of the flapping mass relative to this drag hinge.
On a semi-rigid rotor, the first drag mode of a blade or flapping mass depends on the characteristics not only of the blade or flapping mass but also of the hub. The pulsatance of the first drag mode is then given by the expression:       ω    ⁢          xe2x80x83        ⁢    δ    =            Ω      ⁡              (                                                            e                ·                M                            ⁢                              xe2x80x83                            ⁢              δ                                      I              ⁢                              xe2x80x83                            ⁢              δ                                +                                    k              ⁢                              xe2x80x83                            ⁢              δ                                                      Ω                2                            ⁢              I              ⁢                              xe2x80x83                            ⁢              δ                                      )                    1      /      2      
where kxcex4 is the stiffness of the drag damper fitted between the blade or corresponding flapping mass and the hub of the rotor.
The positioning of the first drag mode of a blade or of the corresponding flapping mass depends upon the modes of the helicopter structure on the ground (fuselage mass, inertia, stiffness of the landing gear and of any tyres which may be fitted to it), these modes of the structure being generally determined by specific tests, adjustment of the first drag mode being obtained by altering the term kxcex4 representing the stiffness of the drag damper.
As a general rule, as the upper limit of the first drag mode xcfx89xcex4, a value close to three-quarters of the nominal rotation frequency xcexa9 of the rotor is taken, so as not to introduce excessively high stresses in the blades of the rotors.
For this reason, when the rotor is started up or stopped, and also at the end of a landing by the helicopter in autorotation, the instantaneous speed of rotation of the rotor intersects the resonance drag frequency situated below the nominal speed. Because of this, and also because of the fairly large range of variations in rotor rotation speeds which are authorised for helicopters in flight, it is necessary to increase damping at the natural vibration frequencies of the blades in drag, and if necessary to reduce this natural frequency by means of drag dampers, which is the reason why these dampers are also termed frequency adapters, the aim being that the blades should be sufficiently damped in drag to avoid going into resonance.
The invention relates more specifically to a drag damper of the general type comprising a tubular damper body in which a piston moving integrally with a damper rod and slidable axially is fitted, and in the damper body delimits two opposing variable volume working chambers, filled with a fluid of which volumes are transferred, by passing the fluid through at least one restriction port arranged between the piston and the body and/or in the piston, between the working chambers when the piston moves in the damper body.
A drag damper of this kind, known in particular from FR 2 063 969, may be hinged on the one hand to a fixed point on the hub or on a bracket connected to the rotor hub, by means connecting the body or one end, external to said body, of the damper rod and on the other to a fixed point on the corresponding flapping mass, at the blade root of this flapping mass or on a device connecting this blade to the hub, by means connecting the end of the rod external to the body or the damper body respectively in the conventional configuration or, in the inter-blade configuration, to two fixed points on two adjacent flapping masses, by means connecting on the one hand the damper body to the fixed point on a flapping mass and on the other, the end of the rod external to the damper body to the fixed point on the other flapping mass.
In the conventional configuration, as the hinging point at one end of the drag damper on the hub or a bracket fixed to the hub is situated between the blade on which the drag damper is hinged at its other end and an adjacent blade, the stiffness of the damper introduces an equivalent angular stiffness, opposed to the angular deflections of the blade relative to the hub about its drag axis. It is thus possible to increase the natural frequency of the blades in drag to escape from the two resonance phenomena mentioned above, with additional damping at the natural drag frequency xcfx89xcex4 of the blade when the phenomena of air and ground resonance occur.
However, it is known these phenomena rarely appear during the life of a helicopter. Most of the time, the drag dampers are subject to forced excitation at the rotation frequency xcexa9 of the rotor, on which the drag dampers dissipate energy to no purpose.
The mean power dissipated in a drag damper of a rotor can be expressed by the following relation: Pd=xcfx80. Kxe2x80x3.f.Xe2, where Kxe2x80x3 is the dissipative stiffness of the damper, f the frequency of the movement applied to the damper (axial movement of the rod-piston assembly in the body) and Xe is the movement of said rod-piston assembly associated with frequency f.
For example, for drag dampers of the type presented above fitted in the conventional configuration to a four-bladed main rotor of a helicopter with a weight of about 8 to 10 tonnes, a comparison of the energy dissipated on the forced excitation at the rotation frequency xcexa9 with that which is dissipated on the natural drag frequency xcfx89xcex4 of the blades gives the following results.
For the same dissipative stiffness Kxe2x80x3 of 400 daN/mm, the forced excitation at xcexa9 with a frequency f of 4.5 Hz corresponds to an associated displacement Xe of 4 mm, that is to say a dissipated power of 900 W, whereas for the natural drag mode xcfx89xcex4 at a frequency f of 2 Hz, causing an associated displacement Xe of 1 mm, there corresponds a dissipated power of about 25 W.
Each drag damper therefore dissipates 97% of its energy on forced excitation at xcexa9. Under these conditions of use, this energy is dissipated to no purpose, which entails substantial fatigue in the components not only of the drag damper but of the means connecting it to the hub and to the flapping mass, and consequently wasted weight due to oversizing of these parts.
On the helicopter with a weight of about 8 to 10 t considered, the forces applied to the drag dampers upon forced excitation at xcexa9 are very high, which may cause incidents in service such as cracks in the yokes connecting the drag dampers to the flapping masses, damage to the drag damper pins at the end where there are connected to the hub and also damage to the fittings connecting the drag dampers to the hub, and rapid deterioration of the ball joints used in these connection devices.
In the inter-blade configuration, the same is true of the yokes, pins and fittings connecting the drag dampers to the corresponding adjacent flapping masses.
To summarise and in other words, the excitation of the drag dampers has a twofold component:
a forced response, at the nominal rotation frequency xcexa9 of the rotor, which adversely affects reliability and the service life of the mechanical parts of the drag damper as well as of its devices connecting it to the hub and to the corresponding flapping mass or masses, and
a natural response, at the natural drag frequency of the corresponding flapping mass, which must be damped to avoid instability.
Known drag dampers of the types presented above damp the unstable mode in two different ways:
by adding stiffness on drag dampers with at least one layer of visco-elastic material, as the loss angle of the visco-elastic material is small, drag dampers of this type being appropriate for helicopters of low and medium tonnage,
by adding damping, when the use of excessively stiff drag dampers would involve excessive penalties in terms of loads, which is suitable for heavy helicopters.
In all cases, the dual-frequency stresses applied to drag dampers cause a substantial loss in the damping of the natural response, at the natural drag frequency of the flapping mass, which is a low frequency relative to the rotation frequency xcexa9 of the rotor, this substantial loss of damping leading to the use of excess damping which adversely affects the reliability of the mechanical parts of the drag dampers as well as of their connections.
The idea underlying the invention is to propose a drag damper for helicopter rotor blades, the damper being of the general type mentioned above and having at least two passages for fluid between the working chambers, each of the passages being dedicated more specifically to dealing with one respectively of the two frequencies corresponding to the forced response at xcexa9 and to the natural response at xcfx89xcex4, which eliminates damping of the dynamic component at the rotation frequency xcexa9 of the rotor, and only provides damping of the drag movement of the blades at the natural drag frequency xcfx89xcex4 with the aim of improving the behaviour in service of all of the components constituting the damper and of the components connecting the damper to the hub and to a flapping mass, or to two adjacent flapping masses, at the same time enabling the weight of the drag damper and of the parts to which it is connected to be reduced.
To this end, the drag damper according to the invention is a drag damper as known from FR 2 063 969, and comprising:
a tubular damper body closed by two end faces,
a main piston slidable axially fitted in the body and delimiting in said body two opposing variable volume working chambers,
a rod moving integrally with the main piston and running substantially axially through at least one end face of the body,
said body and said rod each comprising means of connection to one respectively of the two components between which said drag damper is fitted (the hub and a flapping mass or two adjacent flapping masses),
a fluid filling at least the two working chambers in the body, and
at least one restriction port made in the drag damper, for example in the main piston and/or between the main piston and the body and/or in a part of the body, and capable of restricting fluid passing from one working chamber to the other when the main piston is moved in the body;
wherein the drag damper further comprises at least one channel communicating with the two working chambers, the at least one channel being fitted with a slidable and substantially pressure-tight secondary piston which is subject to loading by an elastic bias tending to return said secondary piston to a neutral position in said channel, with the at least one channel having a greater cross-sectional area than the restriction port, and with the secondary piston and the elastic bias being dimensioned in such a way that the secondary piston has a resonance frequency in the channel substantially equal to the rotation frequency of the rotor, in order to filter a dynamic component of stresses applied to the damper at the rotation frequency of the rotor; and wherein the elastic bias substantially blocks the secondary piston in said channel at the natural drag frequency of the flapping mass; and wherein said at least one restriction port is calibrated so as to damp substantially relative movements of the rod-main piston assembly and of the damper body at a frequency which is substantially equal to the natural drag frequency of the flapping mass.
The drag damper according to the invention has the advantage of comprising at least two passages between the working chambers, one of which, with a larger cross-section that the other, houses a secondary piston, floating and subject to loading from the elastic bias, to filter or suppress the dynamic component at xcexa9 in the stresses, by setting the resonance frequency of the secondary piston in the passage which receives it to the rotation frequency xcexa9 of the rotor, while providing substantial damping for the natural drag mode of the blades at xcfx89xcex4 by means of the other passage with a smaller cross-section, to counter the problems of air and ground resonance, without degrading the filtering of the xcexa9 component.
For an excitation at xcexa9, the resonance of the secondary piston allows practically free circulation of fluid between the working chambers and the channel in which the secondary piston is resonant, since the pressure loss in this channel is very much less than that of the restriction port, through which the flow of fluid is practically nil, while for an excitation at xcfx89xcex4, the elastic bias blocks the secondary piston in the channel which receives it, and the fluid circulates substantially through the restriction port while providing damping, which is applied substantially only at the frequency xcfx89xcex4.
In an advantageously simple embodiment, the channel may be made in the main piston and have longitudinal end portions which each open directly into one respectively of the working chambers.
In this case, according to a basic embodiment, the channel runs substantially axially through the main piston, and the elastic bias advantageously comprises two springs, each running substantially axially through one respectively of the two working chambers and each engaged in one respectively of the longitudinal end portions of said channel, each spring bearing against the secondary piston at one end and at the other end against one respectively of said end faces.
To produce resonance at xcexa9 of the secondary piston and of the fluid displaced by this second piston, this secondary piston and this fluid have an equivalent mass m and the elastic bias has a stiffness K which substantially satisfy the relation K=mxcexa92. This concept of equivalent mass is known, and stems from the fact that the velocity of displacement of fluid in the channel housing the secondary piston is different from the fluid circulation velocities in the working chambers, under the effect of the changes in cross-section. The equivalent mass is that which having a velocity {dot over (x)} equal to that imparted to the piston, would have the same kinetic energy xc2xd m{dot over (x)}2 as that of all of the masses in movement (secondary piston and fluid displaced by this piston).
Consideration of the equivalent mass m means that amplification, if any, of the movement of the fluid with respect to the movement of the secondary piston can be taken into account, in a mode of embodiment of the drag damper utilising the principle of fluid inertia (in which acceleration of the fluid is produced by means of a variation in cross-section of the column of fluid).
According to a variant, the elastic bias may comprise two springs each housed in one respectively of the two longitudinal end parts of said channel, and each bearing at one end against the secondary piston and at the opposite end against the main piston. In this case, to bring about the resonance at xcexa9 of the secondary piston and of the fluid displaced by the latter, this secondary piston and this displaced fluid have an equivalent mass m and the elastic bias has a stiffness K which substantially satisfy the relation       K    =          m      ⁢              xe2x80x83            ⁢                        Ω          2                ⁡                  (                      1            -                          s1              s0                                )                      ,
where s1 and s0 are cross-section areas of the at least one channel and main piston respectively.
According to another variant, precisely utilising the principle of fluid inertia, as mentioned above, the body of the drag damper comprises two tubular casings rigidly connected to each other, one of which is an outer casing closed by said end faces, and the other is an inner casing, arranged inside the outer casing so as to delimit with the latter at least one area with a narrowed cross-section, and in which is fitted the main piston, slidable axially, said inner casing comprising a transverse wall with said rod slidable axially running through it, and in which said channel is made. In this case, it is advantageous that the channel runs substantially axially through said transverse wall of said inner casing, and said elastic bias comprises at least one spring bearing at one end against one of said end faces and at the other against said secondary piston. As a variant, at least one restriction port also runs through said transverse wall of the inner casing or, more generally, a part of the body.
In a simple manner, at least one restriction port may run substantially axially through the main piston, and the latter may be fitted to be slidable and pressure-tight in said body.
Moreover, to achieve substantially the blocking of the secondary piston by the elastic bias and the flow of fluid substantially through the restriction port, in order to obtain high damping at the natural drag frequency xcfx89xcex4 of the flapping mass, the stiffness K of the elastic bias is greater than a threshold value. In the case of singular pressure losses via the restriction port, said threshold value is substantially equal to
xcfx89xcex4.v2.xcex2.s12/s2,
where xcex2 is the coefficient of singular pressure loss of said at least one restriction port; v2 the maximum velocity of circulation of said fluid in said restriction port; and s2 is a cross-sectional area of the at least one channel. The coefficient xcex2 links the pressure loss xcex94P caused by the restriction port and the velocity v2 by the relation: xcex94P=xcex2.v22.
In the case of laminar pressure losses via the restriction port, said threshold value is substantially equal to:       ω    ⁢          xe2x80x83        ⁢          δ      ⁢              xe2x80x83            ·      C        ⁢          xe2x80x83        ⁢          2      ·                        s1          2                s2              ,
where C2 is the laminar pressure loss coefficient linking the pressure loss xcex94P and the velocity v2 by the relation: xcex94P=C2.v2.
To ensure optimum restriction of the fluid at the natural drag frequency xcfx89xcex4, the main piston may have a plurality of restriction port drillings calibrated to different cross sections.
As a variant, at least one restriction port may run through a part of the damper body or said secondary piston.